EP0416460B1 - Hochdruck-Kraftstoffeinspritzeinheit mit Steuerung des Druckes der Kammer für den Spritzzeitpunkt - Google Patents

Hochdruck-Kraftstoffeinspritzeinheit mit Steuerung des Druckes der Kammer für den Spritzzeitpunkt Download PDF

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Publication number
EP0416460B1
EP0416460B1 EP90116619A EP90116619A EP0416460B1 EP 0416460 B1 EP0416460 B1 EP 0416460B1 EP 90116619 A EP90116619 A EP 90116619A EP 90116619 A EP90116619 A EP 90116619A EP 0416460 B1 EP0416460 B1 EP 0416460B1
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EP
European Patent Office
Prior art keywords
valve
timing
plunger
chamber
pressure
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Expired - Lifetime
Application number
EP90116619A
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English (en)
French (fr)
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EP0416460A3 (en
EP0416460A2 (de
Inventor
Timothy A. Warlick
Jeffrey L. Campbell
Lester L. Peters
Michael E. Lucas
Julius P. Perr
Kuang-Wen Thomas Wan
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Cummins Inc
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Cummins Engine Co Inc
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Publication of EP0416460A3 publication Critical patent/EP0416460A3/en
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M57/00—Fuel-injectors combined or associated with other devices
    • F02M57/02—Injectors structurally combined with fuel-injection pumps
    • F02M57/021—Injectors structurally combined with fuel-injection pumps the injector being of valveless type, e.g. the pump piston co-operating with a conical seat of an injection nozzle at the end of the pumping stroke
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M57/00—Fuel-injectors combined or associated with other devices
    • F02M57/02—Injectors structurally combined with fuel-injection pumps
    • F02M57/022—Injectors structurally combined with fuel-injection pumps characterised by the pump drive
    • F02M57/023—Injectors structurally combined with fuel-injection pumps characterised by the pump drive mechanical
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M57/00—Fuel-injectors combined or associated with other devices
    • F02M57/02—Injectors structurally combined with fuel-injection pumps
    • F02M57/022—Injectors structurally combined with fuel-injection pumps characterised by the pump drive
    • F02M57/023—Injectors structurally combined with fuel-injection pumps characterised by the pump drive mechanical
    • F02M57/024—Injectors structurally combined with fuel-injection pumps characterised by the pump drive mechanical with hydraulic link for varying the piston stroke
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M59/00—Pumps specially adapted for fuel-injection and not provided for in groups F02M39/00 -F02M57/00, e.g. rotary cylinder-block type of pumps
    • F02M59/20—Varying fuel delivery in quantity or timing
    • F02M59/30—Varying fuel delivery in quantity or timing with variable-length-stroke pistons

Definitions

  • the present invention relates to unit fuel injectors having an open nozzle and a reciprocating injection plunger that is mechanically actuated by an engine cab shaft. More particularly, the present invention relates to a low speed valve, high pressure unit fuel injector in which the timing metering and the timing chamber pressure are independently controlled.
  • unit fuel injectors having a simplified design have been developed to reduce costs, while providing reliable, precise, and independent control over injector timing and metering.
  • the following patents owned by the assignee of the present application disclose such unit injectors, and are representative of the prior art unit injectors upon which the present invention improves: U.S. 4,471,909, U.S. 4,441,654, U.S. 4,410,138, U.S. 4,410,137. All of these patents disclose fuel injectors having an open nozzle and a reciprocating injection plunger mechanically actuated by an engine camshaft.
  • U.S. 4,721,247 forming the starting point of the invention according to the preamble of claim 1 addresses the problems of achieving high sac pressures throughout the entire range of engine speeds.
  • Perr '247 discloses an open nozzle type unit fuel injector capable of achieving sac pressures exceeding 30,000 psi during injection even at low engine speed.
  • This type of injector is known as a high pressure injector (HPI) and includes a plunger assembly having three plungers arranged to form a hydraulic variable timing fluid chamber between the upper and intermediate plungers and an injection chamber below the lower plunger.
  • HPI high pressure injector
  • An increase in sac pressure is obtained under both low and high speed operating conditions by being designed to achieve high pressures at low engine speed and by being provided with a pressure actuated valve for draining timing fluid from the timing chamber when the engine is operated at higher speed.
  • This patent uses a single spring mounted between the intermediate and lower plunger to bias the intermediate plunger upwardly.
  • the intermediate plunger bias spring also supplies the bias force necessary to operate the pressure actuated relief valve. Accordingly, it becomes very difficult to optimize timing fluid metering without affecting adversely the operation of the pressure actuated relief valve and vice versa.
  • the size of the drain passage from the timing chamber affects both the opening pressure of the pressure limiting valve and the flow rate of timing fluid drained from the timing chamber, through the pressure limiting valve.
  • No known prior art fuel injector incorporates a system for reducing wear, increasing durability, and increasing performance characteristics by means of a variable length hydraulic link forming timing chamber and associated structure whereby the timing fluid metering function can be optimized independently of the pressure limiting valve mechanism which allows high pressure operation of the injector at low engine speed.
  • the prior art fails to disclose an end-of-injection spillport mechanism for maintaining the injector orifices closed to eliminate secondary injection.
  • Another object of the present invention is to achieve the above object with a high pressure unit fuel injector using a dual biasing system in conjunction with the valve in which both biasing devices act on the valve to tend to move the valve toward its closed position, thereby regulating the pressure against which timing fluid must act to control the pressure and discharge of the timing fluid, whereas only one biasing device controls the metering of timing fluid.
  • Still another object of the present invention is to provide a dual spring high pressure three plunger unit fuel injector in which a valve controls timing chamber fluid drainage during injection to limit peak injection pressures while a separate timing fluid spillport located in the injector housing permits a restricted discharge of timing fluid at the end of injection.
  • Timing fluid drain passage is disposed in the intermediate plunger and has a valve seat with an effective cross-sectional flow area that is at least four percent of the coplanar cross-sectional area of the intermediate plunger.
  • the fuel injector includes an injector housing having a central axial bore with a plunger assembly disposed within the central bore.
  • the plunger assembly includes upper, lower, and intermediate plungers.
  • a collapsible timing chamber having a drain passage closable by a valve having a valve element, is formed between the upper and intermediate plungers to receive timing fluid to create a variable length hydraulic link between the upper and intermediate plungers to advance injection timing.
  • valve element is acted upon in part by the upward bias of a timing spring which tends to move the valve element toward its closed position in addition to establishing a biasing force which resists metering of timing fluid into the timing chamber.
  • a timing spring acts directly on the valve element to bias the valve element toward its closed position without affecting the metering of timing fluid.
  • the valve opens to drain timing fluid when the timing fluid pressure exceeds the maximum preset pressure governed by the springs at any time during the injection cycle. Also, after the injection stage is completed by the seating of the lower plunger in the injection chamber portion of the central bore, the valve opens to drain the timing fluid.
  • timing and valve springs enable the biasing force tending to close the valve to be increased, thereby permitting an increased valve area to be exposed to the valve opening pressure without causing this valve area to operate merely as an orifice.
  • This arrangement also obviates in some embodiments the need for a timing spillport due to the increased flow capabilities.
  • the intermediate plunger axial passage for draining timing fluid may be formed having a regulating orifice portion, upstream of the valve seat, with a smaller cross-sectional area than the valve seat. This permits a reduced rate of discharge flow than would otherwise be permitted by the valve seat which governs the operating pressure of the valve. This feature may be incorporated into the valve with or without the dual spring feature.
  • an independently operating timing fluid spillport is used in addition to the pressure relief valve.
  • the valve regulates and limits peak pressures during the injection stroke at high speed and high load conditions as in the above embodiment, but the spillport controls the collapse of the hydraulic link after the injection stroke.
  • the use of the spillport in combination with the valve in a unit fuel injector improves the pressure regulation and performance capabilities of the valve and improves the durability of the valve by reducing the use of the valve at the end of injection. This feature can also be combined with the valve modifications discussed above.
  • the lower edges of the upper plunger portion of the plunger assembly may be formed with straight, nonbeveled edges created by the cylindrical sidewall of the upper plunger intersecting at a perpendicular angle the lower planar surface of the plunger.
  • the cylindrical sidewall passes over the spillport to gradually close the spillport to maintain high spill loads at low speeds and prevent secondary injection.
  • Figure 1 is a cross-sectional view of a dual spring high pressure unit fuel injector according to one embodiment of the present invention.
  • Figure 2a is an enlarged cross-sectional view of the valve mechanism of the fuel injector of Figure 1.
  • Figure 2b is a broken away side view of the valve mechanism of Figure 2a.
  • FIGs 3a and 3b are enlarged broken away cross-sectional views of the valve mechanism employed in the injector of Figures 1, 2a, and 2b wherein the valve mechanism is illustrated in the closed and open positions, respectively.
  • Figures 4a and 4b are views similar to Figures 3a and 3b showing another embodiment of the valve mechanism.
  • FIGS 5a-5d are cross-sectional views of the fuel injector of Figure 1 in the different phases of its operation.
  • Figures 6a and 6b are graphs comparing the operation of the dual spring low speed valve with a prior art single spring low speed valve.
  • FIGS 7a through 7c are three different embodiments of the drain passage which may be formed in the intermediate plunger of an injector designed in accordance with the subject invention.
  • Figure 8 is a graph comparing the sac pressures versus time of the valves of Figures 7b and 7c with the valve of Figures 1-5 at an operating speed of 3,000 rpm.
  • Figure 9 is a graph comparing the sac pressures versus time of the valves of Figures 7b and 7c with the valve of Figures 1-5 at an operating speed of 4,200 rpm.
  • Figure 10 is a graph comparing the highest, average, and lowest pressures, the amount of secondary injection, and the peak-to-peak difference for the valves compared in Figure 8.
  • Figure 11 is a graph comparing the highest, average, and lowest pressures, the amount of secondary injection, and the peak-to-peak difference for the valves compared in Figure 9.
  • Figure 12 is a graph comparing injection duration at various speeds and operating conditions of six test runs for the valves of Figures 7b and 7c and the valve of Figures 1-5.
  • Figure 13 is a cross-sectional view of a dual spring low speed valve high pressure unit fuel injector having a timing chamber spillport according to another embodiment of the present invention.
  • Figures 14a, 14b, and 14c are enlarged cross-sectional views of the timing chamber spillport closure of Figure 13 showing the various stages of spillport closure.
  • Figures 15a, 15b, 15c, and 15d are graphs of the performance of the high pressure unit fuel injector of Figure 13 illustrating the upper plunger travel, camshaft velocity, upper plunger load, and injection pressure versus crank angle.
  • a high pressure unit fuel injector having a pressure limiting valve according to the present invention is shown.
  • the unit fuel injector is of the open nozzle type as shown in commonly assigned U.S. Patent No. 4,721,247 to Perr and is part of a fuel injection system wherein each injector is driven by a rotating camshaft via a conventional drive train assembly (not shown), in which a cam is mounted on a rotatable camshaft and a cam follower rides on the cam to cause the injector plunger to reciprocate in synchronism with camshaft rotation.
  • the fuel injector of Figure 1 is adapted to be used in an injection system including one cam driven unit injector per cylinder and a fuel pump which supplies all the injectors by a common rail or supply line.
  • the fuel injection system requires three common fluid rails (not illustrated) within the cylinder head to communicate with each fuel injector.
  • One rail supplies fuel to each injector for metering into the injection chamber, a second rail drains away fuel that is not injected, and a third supplies timing fluid (which may also be fuel) to vary the timing of the injection event.
  • the effective length of the plunger is caused to increase and advance the beginning of injection, or to decrease and retard the beginning of injection.
  • the fuel pump and engine throttle operate to supply fuel at a variable rail pressure in the first rail, which controls the quantity of fuel injected.
  • the rail pressure may be varied in accordance with pressure/time (PT) metering principles and the timing pressure may be varied in accordance with pressure metering principles as further described in the '247 patent noted above.
  • Figure 1 shows a fuel injector 10 which is intended to be received within a recess contained in the head of an internal combustion engine (not shown).
  • the fuel injector injects a variable quantity of fuel that is metered into the injection chamber 11 (shown collapsed) into the combustion chamber of the engine.
  • the body or housing 16 of the fuel injector is formed of two sections, an injector barrel 12 and a one-piece injector cup 14. Extending axially through the fuel injector is a bore 18 within which a reciprocating plunger assembly 20 is disposed for injecting fuel into the combustion chamber of the internal combustion engine.
  • the plunger assembly 20 is shown in its fully advanced position.
  • the reciprocating plunger assembly 20 includes three plungers.
  • An injection or lower plunger 22 is the lowermost plunger as shown in Figure 1 and injects fuel into the combustion chamber of an engine as discussed below.
  • Serially arranged above lower plunger 22 are an intermediate plunger 24 and an upper plunger 26.
  • a compensating chamber 32 is formed below intermediate plunger 24 and surrounds the upper end of injection plunger 22.
  • a plunger assembly return spring 28 engages the upper end 20 of upper plunger 26 at one end and seats against the top of the injector barrel 12. Return spring 28 biases the upper plunger 26 to return it to an uppermost position within bore 18 as allowed by the injection cam which acts thereon via the drive train assembly.
  • Timing chamber 34 receives hydraulic timing fluid, such as fuel, from a timing fluid passageway 36 formed through the barrel portion 12 of injector housing 16.
  • timing fluid disposed in timing chamber 34 forms a hydraulic link between intermediate plunger 24 and upper plunger 26, and is discharged therefrom under certain conditions through timing chamber drain passage 38 preferably formed centrally axially through intermediate plunger 24.
  • the bottom of timing chamber drain passage 38 opens into compensating chamber 32 and is closed by valve mechanism 40 which is sandwiched between the lower end of intermediate plunger 24 and the upper end of lower plunger 22 disposed within compensating chamber 32.
  • valve mechanism 40 When valve mechanism 40 opens, timing fluid drains from timing chamber 34, through drain passage 38, into compensating chamber 32, and out of the injector through drainage passageway 42. Drainage passageway 42 also may be used for scavenge flow as described below. Valve mechanism 40 controls the pressure of the timing fluid in timing chamber 34, which, in turn, controls the timing of fuel injection as well as the upper limit of injection pressure of the injected fuel.
  • Valve mechanism 40 includes a valve element 44 reciprocally slidable within a valve guide 46, which is an upper portion of lower plunger 22 and has a fluid flow passage 48 ( Figure 3) formed therein.
  • An actuating member such as a cross pin 50 is disposed through a bore 52 in valve body 44 and extends radially outwardly from valve element 44 across most of the width of compensating chamber 32.
  • cross pin 50 may be integrally formed with valve element 44.
  • Cross pin 50 is received without clearance in bore 52 and is disposed, with clearance, through a radial bore 53 formed in lower plunger 22. This bore forms the lower portion of the valve guide 46.
  • Spring guide 54 is disposed beneath the outermost portion of cross pin 50 and around lower plunger 22.
  • Valve mechanism 40 has improved pressure regulation capabilities and improved durability, which goals are accomplished to various degrees by different aspects of this invention.
  • a timing spring 56 which preferably is a coil spring, is positioned within compensating chamber 32 around lower plunger 22.
  • the upper end of timing spring 56 acts against valve mechanism 40 by engaging the outer end of cross pin 50 preferably through the spring guide 54.
  • the lower end of timing spring 56 rests on a seat 57 formed in the bottom of compensating chamber 32.
  • the force of timing spring 56 serves to draw lower plunger 22 upwardly into engagement with intermediate plunger 24 to force the three plungers, lower plunger 22, intermediate plunger 24, and upper plunger 26, together after completion of an injection cycle until metering and timing has commenced for the next cycle.
  • valve spring 60 which also is preferably a coil spring.
  • the lower end of valve spring 60 is seated at the bottom 61 of the hollow bore 58.
  • the upper end of valve spring 60 acts directly against the lower surface of valve element 44.
  • valve spring 60 allows the valve opening pressure to be increased without placing additional unnecessary loads on timing spring 56 which also is required, as discussed above, to bias lower plunger 22 upwardly. This is accomplished in the following manner. Both timing spring 56 and valve spring 60 provide an upward force on the the valve element 44 to maintain the valve in a closed position until the timing fluid pressure exceeds the combined pressure caused by the two springs. Thus, increasing the spring force of valve spring 60 increases the valve opening pressure. However, such an increase does not affect the force tending to resist the metering of timing fluid into the timing chamber which are governed by the timing spring 56.
  • valve spring 60 does not provide a force on the lower plunger 22 relative the injector housing 16 and because the freedom of upward movement of valve element 44 relative to lower plunger 22 is quite limited as discussed below.
  • Pressure regulation is improved using a higher spring force supplied by valve spring 60, and a larger valve area may be used to prevent the valve from operating as an orifice rather than as a timing fluid pressure regulator.
  • valve mechanism 40 has been moved to its closed position to completely seal the timing chamber drain passage 38, any additional biasing force supplied by the valve spring 60 is ineffective to cause additional relative movement.
  • This motion limiting feature is best illustrated in Figures 3 and 4 and is accomplished by limiting the diameter of the radial bore 53 through the lower plunger 22 with a height equal to the thickness of the cross pin 50 passing therethrough plus the desired valve opening distance.
  • the cross pin abuts against the upper surface of the lower plunger radial bore 53. Any further upward expansion of the valve spring 60 or further upward movement of the valve element 44 or the cross pin 50 is prohibited.
  • valve spring 60 In this position, the ends of the valve spring 60 are pushing against spaced portions of the lower plunger 22 which is not formed of separate components relatively movable.
  • the valve spring 60 cannot move any of the components of the plunger assembly 20 or valve mechanism 40 and, therefore, the valve spring 60 does not impair or affect the performance of the timing spring 56.
  • FIG. 5a illustrates the metering and timing stage.
  • Upper plunger 26 has been retracted sufficiently by return spring 28 to uncover timing fluid passageway 36 so that timing fluid enters through the timing fluid passageway 36 into timing chamber 34 and exerts a pressure that separates intermediate plunger 24 from upper plunger 26 by causing timing spring 56 to compress.
  • the amount of separation of upper plunger 26 from intermediate plunger 24 is determined by the equilibrium between the spring force of timing spring 56 and the force produced by the timing fluid pressure acting on the area of intermediate plunger 24. The greater the separation between plungers 24 and 26, the greater the advance of injection timing.
  • Spring 56 also moves plunger 22 upwardly a sufficient extent for fuel to pass into injection chamber 64 adjacent the injection nozzle 66 having a plurality of orifices disposed at the bottom of injection chamber 64. This spring 56 also establishes a biasing force which resists the metering of timing fluid into the timing chamber 34. Then, at the same time that the injection timing is being established by the feeding of timing fluid into timing chamber 34, fuel for injection is caused to flow through a feed orifice of fuel supply passage 62 into the upper portion of injector cup 14.
  • injection chamber 64 During metering of injection fuel, injection chamber 64 will be partially filled with a precisely metered quantity of fuel in accordance with known pressure/time principles whereby the amount of fuel metered is a function of the supply pressure and the total metering time that fuel flows through fuel supply passage 62, which has carefully controlled hydraulic characteristics in order to produce the desired pressure/time metering capability.
  • cam rotation causes upper plunger 26 to be driven downwardly via the drive train assembly.
  • timing fluid is forced back out through timing fluid passageway 36 until the passageway is closed by the leading edge of upper plunger 26.
  • the leading or lower edge of upper plunger 26 may be beveled as is conventional, or it may be straight, forming a right angle with its lower edge to improve closure of a spillport, where used, as described below.
  • the timing fluid becomes trapped between plungers 24 and 26 to form a hydraulic link which causes all three plunger elements to move in unison toward the nozzle tip at the bottom of injection chamber 64.
  • valve mechanism 40 opens to drain timing fluid from timing chamber 34 through timing chamber drain passage 38 and drainage passageway 42 to reduce the pressure to within the preset limits.
  • plunger 22 closes fuel supply passage 62 as it moves downwardly to terminate fuel metering.
  • the fuel metered into injection chamber 64 does not begin to be pressurized until plunger 22 has moved into injection chamber 64 sufficiently to occupy that part of the injection chamber's volume that was not filled with fuel.
  • the distance measured from this point to the point where downward injection plunger travel is completed is termed the "solid fuel height" and determines the point in the plunger's travel when injection actually begins.
  • injector housing 16 may be formed with a timing fluid spillport 70, separate from timing chamber drain passage 38 and drainage passage 42, through which timing fluid drains at the end of the overrun stage.
  • Figure 5d shows the injector scavenge stage after all of the timing fluid has been drained so that plungers 24 and 26 no longer are separated. At this point, the entire injection train, from the injection cam to the nozzle tip, is in solid mechanical contact. In both the overrun and scavenge stages ( Figures 5c and 5d), the system is scavenged of gases and the injector is cooled. In particular, when injection has ended by plunger 22 seating in the nozzle tip, fuel passes from fuel supply passage 62 to an axially relieved portion 63 of lower plunger 22 and travels upwardly into compensating chamber 32 (via a passage not shown) and then out of injector housing 16 via drainage passageway 42. Alternatively, as shown in Figures 1 and 2a, a separate scavenge flow drain port 68 may be used.
  • valve mechanism 40 regulates the pressure of the hydraulic link in timing chamber 34 formed between plungers 24 and 26 to prevent uncontrolled collapse and secondary injection.
  • valve mechanism 40 serves to regulate the pressure in the hydraulic link so that injection is completed at pressures which are close to the preset maximum.
  • valve spring 60 controls the timing fluid pressure independently of timing spring 56, and control of the timing fluid pressure does not affect the setting of the injection timing. This pressure regulating action of valve mechanism 40 also ensures that the duration of injection is minimized and that injection ends sharply, without secondary injection.
  • valve mechanism 40 in its closed and opened positions.
  • valve element 44 moves away from the opening of timing chamber draining passage 38 to permit drainage of timing fluid
  • valve element 44 preferably is less than 0.01 inch away from the passage. That is, the valve element moves less than 0.01 inch between its open and closed positions. Preferably this distance is approximately 0.008 inch.
  • Figure 4 illustrates an alternate embodiment for valve mechanism 40 shown in both the closed and open positions.
  • valve element 44 is lengthened as compared with the first embodiment. Although this increases the mass of the valve element the spring force of the valve spring can be similarly increased to compensate for the increased mass. Additionally, in this embodiment shorter timing spring and valve springs are used to prevent spring buckling.
  • valve body 44 travels 0.356 mm or approximately 0.014 inch between its closed and open positions.
  • the dual spring low speed valve can achieve significantly higher flow areas.
  • the flow area of the valve seat formed by the passage adjacent the valve is approximately 1.5% of the coplanar cross-sectional area of the intermediate plunger.
  • the area of the valve seat formed by passage 38 adjacent valve element 44 is approximately 4.4% of the coplanar cross-sectional area of intermediate plunger 24.
  • the area of the passageway in the present invention is almost three times the area of the prior art passageways. The increased area is made possible by using two separate springs 56 and 60 and produces the numerous improvements and advantages achieved by the present invention.
  • the dual spring unit fuel injector of the present invention improves the pressure regulation by the valve and increases the flow area capability of the valve to eliminate the need in some circumstances for a spillport in the injector housing.
  • the increased spring load achieved by the valve spring (from 12.1 lbs to 23.6 lbs in one embodiment) enables the cross-sectional flow area of the timing chamber drain passage (the valve seat area) to be increased while still maintaining the same valve opening pressure (because opening pressure equals spring load divided by timing chamber drain passage area).
  • the timing chamber drain passage is large enough that the main restriction in the valve is the distance the valve body moves (the valve opening size) during opening of the valve mechanism, thereby improving the regulating capability.
  • the combination of the mass of the valve mechanism combined with one third of the mass of the timing and valve springs is less than the mass of equivalent components in single spring type injectors. This decreases the inertia and provides a better and quicker valve response. This allows the valve mechanism to open and close more frequently during a given time period to act as a regulator rather than an orifice, and to decrease the possibility of secondary injection caused by a relatively slow valve response. Additionally, at rated speed (e.g., 5,000 rpm) and load, the number of large oscillations of the valve is decreased and the pressure regulation capabilities are increased over three-hole valve designs, as shown graphically in Figure 6. The decrease in the number of valve oscillations improves the valve and spring durability because a single spring tends to fatigue and wear faster than a dual spring design. This valve design allows the spillport to be eliminated obtaining better performance than a three hole valve and spillport combination design.
  • timing chamber drain passage 38 of at least two portions having different cross-sectional areas.
  • the remainder of the fuel injector is as described above.
  • Figure 7a illustrates, for comparison, the drainage passage of the embodiments of Figures 1-5.
  • Figures 7b and 7c are two different versions of the alternate embodiment having a multiple area drain passage 38.
  • a main orifice is formed at the bottom of timing chamber drain passage 38 adjacent valve element 44.
  • the main orifice or valve seat 38a has a cross-sectional area that is selected to control the opening operating pressure of valve mechanism 40, and therefore the pressure of timing fluid within the timing chamber 34.
  • Valve seat 38a controls the valve opening for a given spring rate and preload.
  • valve seat 38a controls the injection pressure of the fuel injector.
  • Regulating orifice 38b is formed upstream of valve seat 38a and has a cross-sectional area that is smaller than the cross-sectional area of valve seat 38a.
  • Regulating orifice 38b in combination with valve seat 38a and the valve opening distance, controls the flow rate of timing fluid through timing chamber drain passage 38 in intermediate plunger 24.
  • the cross-sectional area of regulating orifice 38b is smaller than that of valve seat 38a, the flow rate is dominated by the size of regulating orifice 38b and the flow rate is less than otherwise permitted by the size of the valve seat.
  • the opening pressure of valve mechanism 40 remains the same.
  • the effective flow area is reduced and controlled by the regulating orifice.
  • the injection pressure can be more easily regulated to achieve better injection characteristics. This provides for a smoother discharge flow and prevents an undesirably large discharge flow which could result in a large pressure drop before the valve mechanism closes.
  • the exiting timing fluid pressure remains higher after the valve mechanism is open. This increased pressure and the lower flow rate prevent the hydraulic link from totally collapsing before the lower plunger 22 is seated at the bottom of the injection chamber 64 and maintains the lower plunger in its seated position to prevent secondary injection.
  • valve element 44 The restriction imposed on the flow by regulating orifice 38b when combined with the inertia effect of valve element 44 creates high peak pressures which can be countered by using a lower spring load.
  • the lower spring load improves the durability and increases the life of the system. Additionally, the improved pressure regulation accomplished with this design typically yields higher mean injection pressures, lower peak-to-peak values, and shorter injection duration. These benefits appear to increase with increasing operating speeds.
  • regulating orifice housing 39 is formed as an insert portion disposed within timing chamber drain passage 38.
  • a drain passage 38 having a predetermined main orifice or valve seat cross-sectional area any one of a plurality of different regulating orifice housings 39 may be used, each one having a regulating orifice 38b having a different cross-sectional area.
  • a regulating orifice housing 39 having regulating orifice 38b may be selected so that the area of regulating orifice 38b creates the desired timing fluid discharge flow rate.
  • an intermediate plunger 24 having a predetermined regulating orifice 38b area is shown.
  • valve seat 38a and regulating orifice 38b differ, by varying these orifices, one of the operating pressure and the timing fluid discharge flow rate can be changed without altering the operating characteristics of the other.
  • Figures 8-12 this multiple area orifice design achieves better pressure regulation, higher mean injection pressures, shorter injection duration, reduced spring stress, and the elimination of some secondary injections.
  • Figures 8 and 9 compare the performance of the valve having the regulating orifice of Figures 7b and 7c with the dual spring valves of Figures 1-5 by comparing the attained sac pressures during injection at operating speeds of 3,000 rpm and 4,200 rpm. Note that secondary injection is reduced with the modified valve of Figures 7b and 7c.
  • Figures 10 and 11 are bar graphs comparing various characteristics of the two types of valve systems as derived from the graphs of Figures 8 and 9.
  • Figures 10 and 11 compare high, average, and low sac pressures, incidence of secondary injection, and differences between high peak and low peak sac pressures.
  • Figure 12 compares six test runs of the valve having the regulating orifice with a valve having no regulating orifice, three at 3000 rpm and three at 4200 rpm at three different operating conditions indicated as A, B, and C. Due to the difficulty in metering a precise predetermined quantity of fuel, the duration is determined for a given amount of fuel to render valid comparisons.
  • valve with the regulating orifice achieves larger high pressures, larger low pressures, and larger average pressures at both operating speeds, while significantly reducing occurrences of secondary injection and reducing the difference between the peak-to-peak (highest-to-lowest) pressures.
  • timing fluid discharge passage having a main orifice or valve seat and a reduced area upstream regulating orifice may be used without the dual spring system for the valve mechanism.
  • this timing fluid discharge orifice may be used with HPI fuel injectors without the valve disclosed herein (such as those of the '247 patent) as well as with other fuel systems using differently operating fuel injectors to control the injection pressure directly.
  • FIG. 13 illustrates an alternative embodiment of the fuel injector of the present invention in which two additional important features are shown.
  • the first feature involves forming a timing chamber spillport 70 in barrel 12 of injector housing 16 in addition to valve mechanism 40 with its timing chamber drain passage 38 and drainage passageway 42.
  • the spillport 70 is used to drain timing fluid from timing chamber 34 during the overrun stage of the injection cycle after the fuel has been injected into the engine cylinders.
  • overrun begins when injection ends by the tip of lower plunger 22 contacting its seat in the nozzle tip.
  • upper plunger 26 continues to move downwardly, timing fluid is forced out of timing chamber 34 and the hydraulic link therein collapses.
  • the timing fluid drains through timing chamber spillport 70.
  • timing chamber spillport 70 and the valve means controlled passage 38 independently control the drainage of timing fluid and the two fuel draining paths operate separately and at different times of the injection cycle.
  • the valve mechanism operates to control pressure in the timing chamber 34 by regulating and limiting peak timing fluid pressure during the second and the third stages, the injection and overrun stages, particularly at high speed and high load operating conditions, while the spillport 70 controls collapse of the hydraulic link after injection during the overrun stage.
  • timing chamber spillport 70 decreases the use of valve mechanism 40 by over 50% as the valve mechanism will operate only when timing fluid pressure exceeds a preset limit and not during the overrun stage of each injection cycle.
  • valve mechanism 40 would operate only during high load and middle to high speed conditions, because in other operating modes the timing fluid pressure typically does not approach the level set by the valve mechanism. This extends the life and improves the durability of valve mechanism 40 and particularly of valve spring 60 and the valve seat areas which are highly stressed. Additionally, the overall injection performance may be improved by this configuration because spillport induced hydraulic link collapse and valve induced hydraulic link collapse can be separately set and optimized.
  • the spillport is sized to control collapse of the timing chamber 34 after injection has ended and the valve mechanism is selected and set to limit peak injection pressures during injection.
  • valve mechanism 40 can operate over a much smaller pressure range which improves the quality of the valve regulation.
  • the timing chamber spillport 70 has been combined with valves using a dual spring system in high pressure fuel injectors.
  • the same advantages are obtainable by combining the timing chamber spillport 70 with any type of valve. Even without the dual spring configuration or the dual orifice timing fluid discharge passage 38, using a spillport 70 with a valve mechanism improves the pressure regulation capabilities of the valve mechanism and therefore the fuel injector 10.
  • the second modification of the high pressure unit fuel injector with timing chamber pressure control which is shown in Figure 13 includes providing an improved closure formed on the lower portion of upper plunger 26 for timing chamber spillport 70.
  • Upper plunger 26 is formed of a cylindrical sidewall 72 and a planar lower wall 74.
  • Sidewall 72 intersects lower wall 74 at a generally sharp perpendicular angle in the vicinity of timing chamber spillport 70, and this perpendicular relationship may extend completely around upper plunger 26.
  • sidewall 72 extends parallel to the inner walls of axial bore 18 all the way to lower wall 74. There is no beveled or chamfered portion.
  • FIG 13 which illustrates the end of injection stage of operation similar to that of Figure 5c in which timing chamber 34 is being collapsed
  • camshaft velocity and upper plunger velocity are very low
  • the lower wall 74 of upper plunger 26 is nearing direct mechanical contact with the upper wall of intermediate plunger 24.
  • the nonbeveled portion of sidewall 72 serves to close timing chamber spillport 70 at least partially but possibly completely. This decreases the effective area of timing chamber spillport 70 at the end of injection notwithstanding the decreasing plunger speed to maintain a relatively high pressure in the timing chamber 34 which, in turn, maintains a sufficiently high spill load on lower plunger 22 to prevent lower plunger 22 from rising off of its seat in injection chamber 64 and causing secondary injection. Additionally, the closure of spillport 70 in this manner has been found to increase power and reduce unburned hydrocarbon emissions in the engine.
  • the bevel is formed to eliminate flow restrictions during the metering of timing fluid into the timing chamber.
  • the use of a non-beveled upper plunger 26 also eliminates these flow restrictions by either using a larger timing fluid metering port, opening the metering port further, or forming an undercut or internal barrel groove in the barrel at the location of the metering port. All of these solutions prevent flow restrictions during filling while maintaining high spill loads to prevent secondary injections during spilling.
  • the preferred solution is to increase the size of the metering port and to open the metering port further.
  • Figure 15 is a series of four graphs, plotting the upper plunger travel, the camshaft velocity, the upper plunger load, and the injection pressure versus crank angle. Corresponding crank angles on each graph are so indicated.
  • the upper plunger travel is shown.
  • cam velocity is shown.
  • the standard HPI unit fuel injector has a relatively low load on the upper plunger near the end of injection as shown by the dip identified as "load on standard HPI" illustrated in Figure 15c. As shown in Figure 15d, this low load permits the injection or lower plunger to rise off of its seat and create secondary injection.
  • using the spillport closure on the upper plunger of the present invention increases the upper plunger load as compared with known high pressure injector loads as shown in broken line in Figure 15c.
  • timing spillport closure is not limited to high pressure fuel injectors using low speed valves but may be used with any fuel injector having a timing chamber spillport.
  • the high pressure unit fuel injector provides improved pressure regulation with increased fluid flow capabilities in the valve. Because two separate springs are used with the valve mechanism, the timing can be optimized simultaneously with and separately from setting the required operating pressure of the valve mechanism. The increased flow capabilities therefore can be achieved while maintaining the valve opening pressure.
  • the use of two springs also improves the durability of the valve mechanism and the injector as a whole by spreading the spring loads over two springs to reduce spring fatigue. These advantages are significant improvements over prior single spring systems in which one spring was used to both control the valve pressure and control timing by biasing the lower plunger.
  • the valve opening pressure and the discharge flow rate can be controlled separately by varying the area of the draining passage in the intermediate plunger to further improve and optimize operation of the fuel injector.
  • a timing chamber spillport may be provided to drain timing fluid after injection and to supplement the valve and improve operation of the valve, and an improved closure for the timing chamber spillport may be used to prevent secondary injections.
  • the high pressure unit fuel injector of the present invention finds application in a large variety of internal combustion engines.
  • One particularly important application is for small compression ignition engines adopted for automotive uses such as powering automobiles.
  • Lighter truck engines and medium range horsepower engines also could benefit from the use of fuel injectors according to the present invention.

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  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Fuel-Injection Apparatus (AREA)

Claims (12)

  1. Hochdruck-Kraftstoffeinspritzeinheit zur Kraftstoffeinspritzung in die Verbrennungskammer eines Verbrennungsmotors wobei die Einspritzeinheit ein oberes und ein unteres Ende aufweist,
    mit einem Einspritzgehäuse (16), das eine zentrale Bohrung (18) und am unteren Ende eine Einspritzöffnung (66) aufweist, die zentrale Bohrung (18) mit der Verbrennungskammer verbindet,
    mit einem oberen Kolben (26) und einem unteren Kolben (22), die für eine Hin- und Herbewegung in der zentralen Bohrung (18) angeordnet sind,
    mit einer hydraulischen Synchronisationsvorrichtung zur Veränderung der Synchronisation der Einspritzung einer abgemessenen Kraftstoffmenge, wobei die Synchronisationsvorrichtung einen mittleren Kolben (24) aufweist, der zwischen dem oberen Kolben (26) und dem unteren Kolben (22) für eine Hin- und Herbewegung in der zentralen Bohrung (18) angeordnet ist, und eine zwischen dem oberen Kolben (26) und dem mittleren Kolben (24) angeordnete, zusammendrückbare Synchronisationskammer (34) bildet,
    mit einer in dem Zwischenkolben (24) angeordneten Abflußleitung (38) zum Ableiten der Synchronisationsflüssigkeit aus der Synchronisationskammer (34),
    mit einer von einer geschlossenen in eine geöffnete Position bewegbaren Ventilvorrichtung (40) zur Begrenzung des Druckes der Synchronisationsflüssigkeit in der Synchronisationskammer (34) durch Ablassen der Synchronisationsflüssigkeit aus der Synchronisationskammer (34) durch die Abflußleitung (38) und
    mit einer Vorspannvorrichtung, die für eine Steuerung des Ventilöffnungsdruckes auf die Ventilvorrichtung (40) einwirkt, wobei die Vorspannvorrichtung ein erstes Vorspannelement (56) aufweist, um eine Vorspannkraft zu erzeugen, die das Einmessen von Synchronisationsflüssigkeit in die Synchronisationskammer (34) begrenzt und die die Ventilvorrichtung (40) in Richtung ihrer geschlossenen Position zu bewegen drängt,
    dadurch gekennzeichnet,
    daß das erste Vorspannelement (56) so angeordnet ist, daß es den unteren Kolben (22) derart vorspannt, daß der untere Kolben (22) im Eingriff mit dem mittleren Kolben (24) steht und
    daß ein zweites Vorspannelement (60) vorgesehen ist, um eine zusätzliche Vorspankraft hinzuzufügen, die die Ventilvorrichtung (40) in Richtung ihrer geschlossenen Position zu bewegen drängt, ohne die Vorspannkraft zu erhöhen, die dem Einmessen der Synchronisationsflüssigkeit in die Synchronisationskammer (34) entgegenwirkt.
  2. Hochdruck-Kraftstoffeinspritzeinheit nach Anspruch 1, wobei eine Vorrichtung vorgesehen ist, die die Bewegung der Ventilvorrichtung (40) in Richtung ihrer durch das zweite Vorspannelement (60) vorgegebenen geschlossenen Position begrenzt, und wobei vorzugsweise, das erste Vorspannelement (56) und das zweite Vorspannelement (60) jeweils eine Feder aufweisen, das erste Vorspannelement (56) um den unteren Kolben (22) herum angeordnet ist und das zweite Vorspannelement (60) innerhalb eines im oberen Ende des unteren Kolbens (22) angeordneten Hohlbereichs (58) angeordnet ist.
  3. Hochdruck-Kraftstoffeinspritzeinheit nach einem der vorangegangenen Ansprüche, wobei die Ventilvorrichtung (40) ein Ventilelement (44) aufweist, das eine obere Fläche aufweist, die abdichtend an der unteren Öffnung der Abflußleitung (38) anliegt, wenn die Ventilvorrichtung (40) in ihrer geschlossenen Position ist.
  4. Hochdruck-Kraftstoffeinspritzeinheit nach Anspruch 2 und, vorzugsweise, nach Anspruch 3, wobei der Hohlbereich (58) im oberen Ende des unteren Kolbens (22) als eine Ventilführung (46) ausgebildet und das Ventilelement (44) in dieser Ventilführung (46) verschiebbar aufgenommen ist und wobei das zweite Vorspannelement (60) eine direkt nach oben gerichtete Vorspannkraft auf die untere Fläche des Ventilelementes (44) ausübt.
  5. Hochdruck-Kraftstoffeinspritzeinheit nach Anspruch 4, wobei die Ventilführung (46) mit neben der Ventilführung (46) angeordneten axial ausgerichteten Öffnungen (53) ausgestaltet ist und wobei das Ventilelement (44) einen Stift (50) aufweist, der an dem Ventilelement (44) befestigt ist und in entgegengesetzten Richtungen in die entsprechenden Öffnungen (53) hineinragt, um die Grenzen der Bewegung des Ventilelements (44) relativ zu dem unteren Kolben (22) festzulegen, und wobei, vorzugsweise der Abstand, über den sich das Ventilelement (44) zwischen seiner offenen und seiner geschlossenen Position hin- und herbewegt, ungefähr 0.025 cm (0.01 inch) beträgt.
  6. Hochdruck-Kraftstoffeinspritzeinheit nach einem der vorangegangenen Ansprüche, wobei das Ventilelement (44) eine geringe Masse aufweist, um die Trägheitseffekte bei der Bewegung des Ventilelements (40) zu begrenzen und um die Ansprechzeit des Ventilelements (44) zu verringern.
  7. Hochdruck-Kraftstoffeinspritzeinheit nach einem der vorangegangenen Ansprüche, wobei die Abflußleitung (38) innerhalb des mittleren Kolbens (24) angeordnet ist und auf der unteren Fläche des mittleren Kolbens (24) einen Ventilsitz (38a) für das Ventilelement (44) bildet, wobei der Durchflußquerschnitt des Ventilsitzes (38a) mindestens vier Prozent des in derselben Ebene liegenden Querschnittes des mittleren Kolbens (24) beträgt.
  8. Hochdruck-Kraftstoffeinspritzeinheit nach einem der vorangegangenen Ansprüche,
    wobei die Abflußleitung (38) einen nächst dem Ventilelement (44) angeordneten Ventilsitz (38a) aufweist,
    wobei der Ventilsitz (38a) die effektive Querschnittsfläche auf dem Ventilelement (44) bestimmt, auf die der Druck der Flüssigkeit in der Synchronisationskammer (34) einwirkt, während das Ventilelement (44) geschlossen ist,
    wobei die Abflußleitung (38) eine vom Ventilsitz (38a) strömungsaufwärts angeordnete Regulieröffnung (38b) aufweist,
    wobei die Regulieröffnung (38b) eine effektive Querschnittsfläche aufweist, die kleiner als die effektive Querschnittsfläche des Ventilsitzes (38a) ist, wobei der Ventilsitz (38a) den Öffnungsdruck des Ventilelements (44) steuert, und
    wobei die Regulieröffnung (38b) die Ausflußrate der Synchronisationsflüssigkeit steuert.
  9. Hochdruck-Kraftstoffeinspritzeinheit nach einem der vorangegangenen Ansprüche,
    wobei eine Synchronisationskammer-Absteuerbohrung (70) in dem Einspritzgehäuse (12) ausgebildet ist, um die Synchronisationsflüssigkeit von der Synchronisationskammer (34) abzuleiten,
    wobei die Synchronisationskammer-Absteuerbohrung (70) so angeordnet ist, daß sie nur dann geöffnet ist, wenn der untere Kolben (22) sich seiner untersten Position nähert, in der die Einspritzöffnung (66) geschlossen ist,
    wobei die Synchronisationskammer-Absteuerbohrung (70) eine genügend große Durchflußratenkapazität aufweist, um die Synchronisationsflüssigkeit aus der Synchronisationskammer (34) vorrangig durch die Absteuerbohrung (70) und weniger über das Ventilelement (40) abzuleiten, wenn sich der obere Kolben (26) seiner untersten Position nähert.
  10. Hochdruck-Kraftstoffeinspritzeinheit nach Anspruch 9, wobei die Synchronisationskammer-Absteuerbohrung (70) so dimensioniert ist, daß während der Annäherung des unteren Kolbens (22) an seine unterste Position der Abfluß der Synchronisationsflüssigkeit so begrenzt ist, daß ein genügend hoher Druck auf dem unteren Kolben (22) aufrecht gehalten wird, um den unteren Kolben (2) in seiner untersten Position zu halten.
  11. Hochdruck-Kraftstoffeinspritzeinheit nach Anspruch 9 oder 10,
    wobei der obere Kolben (26) eine zylindrische Seitenwand (72) und eine im wesentlichen ebene untere Oberfläche (74) aufweist,
    wobei die zylindrische Seitenwand (72) die untere Fläche (74) in einem im wesentlichen scharfen rechten Winkel schneidet, so daß die Synchronisationskammer-Absteuerbohrung (70) kontinuierlich nach und nach schrittweise durch die zylindrische Seitenwand (72) des oberen Kolbens (26) abgedeckt wird, wenn sich der obere Kolben (26) seiner untersten Position nähert und sich die Synchronisationskammer (34) ihrer im wesentlichen zusammengedrückten Einstellung nähert.
  12. Hochdruck-Kraftstoffeinspritzeinheit nach Anspruch 11, wobei die zylindrische Seitenwand (72) derart ausgebildet ist, daß sie zumindest teilweise die Synchronisationskammer-Absteuerbohrung (70) verschließt, wenn sich der obere Kolben (26) seiner untersten Position nähert und wenn sich die Geschwindigkeit des oberen Kolbens (26) verringert, wobei die effektive Fläche der Synchronisationskammer-Absteuerbohrung (70) verringert wird, um in der Synchronisationskammer (34) den hohen Flüssigkeitsdruck aufrechtzuhalten, wodurch wiederum der hohe, abwärts gerichtete Druck auf den unteren Kolben (22) aufrechterhalten wird, um eine sekundäre Einspritzung zu verhindern.
EP90116619A 1989-09-05 1990-08-30 Hochdruck-Kraftstoffeinspritzeinheit mit Steuerung des Druckes der Kammer für den Spritzzeitpunkt Expired - Lifetime EP0416460B1 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
US07/402,893 US4986472A (en) 1989-09-05 1989-09-05 High pressure unit fuel injector with timing chamber pressure control
US402893 2003-03-28

Publications (3)

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EP0416460A2 EP0416460A2 (de) 1991-03-13
EP0416460A3 EP0416460A3 (en) 1991-09-25
EP0416460B1 true EP0416460B1 (de) 1994-10-12

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US (1) US4986472A (de)
EP (1) EP0416460B1 (de)
JP (1) JP2524657B2 (de)
DE (1) DE69013283T2 (de)

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US5209403A (en) * 1991-07-12 1993-05-11 Cummins Engine Company, Inc. High pressure unit fuel injector with timing chamber pressure control
US5301876A (en) * 1991-07-12 1994-04-12 Cummins Engine Company, Inc. Unit injector with interchangeable subassembly for converting from open nozzle to closed nozzle operation
EP0529630B1 (de) * 1991-08-30 1996-03-27 Nippondenso Co., Ltd. Brennstoffeinspritzvorrichtung für Brennkraftmaschine
US5323964A (en) * 1992-03-31 1994-06-28 Cummins Engine Company, Inc. High pressure unit fuel injector having variable effective spill area
US5275337A (en) * 1992-06-15 1994-01-04 Cummins Engine Company, Inc. High pressure limiting valve with fast response and fuel injector equipped therewith
US5299738A (en) * 1992-09-16 1994-04-05 Cummins Engine Company, Inc. High pressure fuel injector with cushioned plunger stop
US5277162A (en) * 1993-01-22 1994-01-11 Cummins Engine Company, Inc. Infinitely variable hydromechanical timing control
US5423301A (en) * 1994-02-17 1995-06-13 Cummins Engine Company, Inc. Timing control valve for hydromechanical fuel system
US5487508A (en) * 1994-03-31 1996-01-30 Caterpillar Inc. Injection rate shaping control ported check stop for a fuel injection nozzle
JPH0861181A (ja) * 1994-08-25 1996-03-05 Mitsubishi Electric Corp 燃料噴射装置
DE19528807C2 (de) * 1995-08-05 2000-06-08 Dahlmann Gerd Uwe Kraftstoff-Einspritzventil mit integrierter Pumpe
US5713335A (en) * 1995-09-12 1998-02-03 Cummins Engine Company, Inc. Variable injection timing and injection pressure control arrangement
US5666913A (en) * 1996-05-29 1997-09-16 Cummins Engine Company, Inc. Variable timing cam follower lever assembly
JPH10122090A (ja) 1996-10-14 1998-05-12 Komatsu Ltd ディーゼルエンジンの燃料噴射装置
US5778841A (en) 1997-02-26 1998-07-14 Cummins Engine Company, Inc. Camshaft for internal combustion engines
US5894991A (en) * 1997-08-22 1999-04-20 Cummins Engine Company, Inc. Unit injector with hard stop timing plunger
DE10006111A1 (de) * 2000-02-11 2001-08-30 Bosch Gmbh Robert Kraftstoffeinspritzventil
DE10134529C2 (de) * 2001-07-16 2003-07-03 Bosch Gmbh Robert Injektor mit hintereinandergeschalteten, nach innen öffnenden Ventilen
US7021558B2 (en) * 2003-04-25 2006-04-04 Cummins Inc. Fuel injector having a cooled lower nozzle body
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Also Published As

Publication number Publication date
EP0416460A3 (en) 1991-09-25
EP0416460A2 (de) 1991-03-13
JPH03206351A (ja) 1991-09-09
DE69013283D1 (de) 1994-11-17
US4986472A (en) 1991-01-22
JP2524657B2 (ja) 1996-08-14
DE69013283T2 (de) 1995-02-16

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